Hydraulic pump or motor



Oct. 12, 1965 v. BUSH ETAL. 3,211,105

HYDRAULIC PUMP OR MOTOR Original Filed Jan. 11. 1962 3 Sheets-Sheet 1 M/Z 6/ K C5;

/A VE/i/ 7019f A 7 7 arn'eg Oct. 12, 1965 v. BUSH ETAL 3,211,105

HYDRAULIC PUMP OR MOTOR Original Filed Jan. 11, 1962 3 Sheets-Sheet 2\I. BUSH ETAL HYDRAULIC PUMP OR MOTOR Original Filed Jan. 11, 1962 3Sheets-Sheet 3 lA/VENTOPI ,4 r m y United States Patent Originalapplication Jan. 11, 1962, Ser. No. 165,685. Divided and thisapplication Jan. 6, 1965, Ser. No.

4 Claims. (Cl. 103161) This application is a division of applicationSerial No. 165,685, filed January 11, 1962, relating to hydraulic pumpor motor units of the type having radially disposed fluid chambers eachdefined by a mating piston and cylinder intermediate to opposingreaction elements.

A major problem of any hydraulic pump or motor unit is film lubricationbetween adjacent mating members adapted to slide along one another,while simultaneously being pressed together with loads of largemagnitude. The load supporting capacity of the film lubrication must beadequate to eliminate metal-to-metal contact at the maximum load and atany intermediate loads. Similarly, throughout all outputs and speeds ofthe unit, the power consumption caused by leakage or friction of thefilm must be maintained at a tolerable minimum. The success of the filmlubrication, however, depends on a second problem, the problem ofwithstanding cocking forces or couples applied between the matingmembers.

A couple between two members separated from each other by a fluid filmcauses tilting of the members relative to one another until balanced bya resisting couple of equal magnitude in the opposite direction. Thisresisting couple is developed initially by a shift asymmetrically of thefilm thickness and pressures, causing localized pressure concentrationsof the film. These localized pressure concentrations upon exceeding themaximum allowable film pressures, degenerate to localized metal-to-metalcontacts. Thereafter any deficinecy of the film support is made up bythe direct metal-to-metal support. Consequently, any couple betweenmating members that causes localized pressure concentrations within theseparating film to exceed the maximum allowable film pressure isextremely harmful to the unit and greatly shortens the operating life ofthe unit.

In a piston-type unit, the fluid pressure confined by the mating pistonand cylinder exert an outward force on the piston in the direction ofthe longitudinal axis of the cylinder. The outward force is absorbed byspaced reaction elements opposing the piston and cylinder. The reactionelements are moved toward and away from one another by appropriatestructure to transform between the energy of fluid pressure and work. Ashaft member generally is associated with one of the members to harnessthe Work in the form of rotating shaft torques. The shaft torque must beopposed by an equal and opposite reaction torque through the housingstructure of the unit, which includes the reaction elements.

It is commonplace in existing piston-type units that the pressure forceacting along the longitudinal axis of the cylinder always extendsthrough the center of the shaft. A torque about a point is representedas a force acting in a direction at some normal distance or moment armfrom the point, and is the product of the two. When the pressure forceextends through the center of the shaft, it

thus has no direct moment arm about the shaft. Thus the only waytransformation between fluid pressure forces and shaft torque can occuris by indirectly applied force such as a couple between the adjacentmating members. Since this couple must equal the shaft torque and thusbe of large magnitude, film support fails causing metal-tometal contact.

3,211,105 Patented Oct. 12, 1965 It is also commonplace in conventionalradial piston hydraulic units that the shaft is directly connected tothe cylinder or member rotating on one of the reaction elements. Theother reaction element is mounted for free rotation about an axis spacedfrom the shaft axis. Rotation of the shaft only indirectly, through theinteraction of the other members including the piston and cylinder,causes the movement of the reaction elements and thus reciprocation ofeach piston and cylinder. This further requires that couples be broughtin play to transform between the pressure forces and shaft torque.

There is further a greatly increasing demand to use hydraulic equipmentat pressures exceeding 5,000 .p.s.i. and flow rates approaching 200g.p.m. Also, to reduce the ratio of overall unit weight to outputhorsepower, operating shaft speeds of 5,000 rpm. or higher have beentried and achieved in some commercial embodiments. However, under thesesevere operating conditions, the aforementioned problems relating tococking of, and film support between, adjacent mating members becomeincreasingly complex and important, and more often the direct cause offailure.

The bearing structure supporting the various rotating members must becapable of withstanding the large radial loads applied to the memberswithout excessive torque loss because of friction. Cantileverapplication of these radial loads causes slight transverse deflectionand/ or axial misalignment of the shaft. Thus self-aligning bearingmeans capable of compensating these tendencies must be used;particularly in a film lubrication bearing where mating surfaces mustremain parallel.

Another practical drawback of existing commercial hydraulic pump ormotor units is complex and expensive adjusting structure commonly usedfor varying the fluid flow per cycle of the unit. The limitations ofcommercial manufacturing plus the continuing battle to reduce costsrequire an easily produced structure capable of accurate and dependableoperation.

Accordingly, an object of this invention is to provide a hydraulic pumpor motor unit of the radial piston and cylinder type with improveddisposition of the various members relative to each other to eliminatereaction couples between adjacent members caused directly by theconfined fluid pressures.

Another object of this invention is to provide a hydraulic pump or motorunit having fluid film lubrication between all adjacent mating surfacesadopted to slide along one another while simultaneously being forcedtogether by a large load, the fluid film lubrication being pressurizedfrom a high pressure source through a fluid restrictor to establish andmaintain a balanced film at all operating conditions of the unit.

Another object of this invention is to provide a hydraulic film bearingcapable of supporting a rotating shaft member subjected to a largeradial load while accommodating moderate angular misalignment of thelongitudinal shaft axis from its normal axial position.

Another object of this invention is to provide a hydraulic pump or motorunit having simple economical structure including a double eccentricmounting operable to varying the fluid flow per cycle of theunit.

These and other objects will be more fully appreciated after a completedisclosure of the subject invention given in the following specificationand the accompanying drawings forming a part thereof, wherein:

FIG. 1 is a longitudinal section view of the subject invention asembodied in a hydraulic pump or motor unit, the view being takengenerally from line 11 of FIG. 2;

FIG. 2 is a front section view of the subject invention as takengenerally from line 2-2 of FIG. 1;

FIG. 3 is a side elevational view, similar to that shown s.) in FIG. 1,of a pintle-crank member used in the subject invention;

FIG. 4 is a section view as taken from line 44 of FIG. 3;

FIG. 5 is a section view as taken from line 55 of FIG. 3;

FIG. 6 is an enlarged section view as taken from line 6-6 of FIG. 3, thefigure including operating fluid pressure forces and distributions onthe member;

FIG. 7 is a front elevational view, partly shown in section from line 77of FIG. 8 and of reduced size compared thereto, of a self-aligning filmbearing unit of the subject invention;

FIG. 8 is an enlarged section view as taken generally from 'line 88 ofFIG. 7;

FIG. 9 is a diagrammatic representation of the adjusting structure ofthe subject invention, the view being a greatly enlarged elevationalview similar to FIG. 2; and

FIG. 10 is a longitudinal center section view of a piston used in thesubject invention.

Referring now to FIGS. 1 and 2 of the drawings, the disclosed hydraulicunit 10 includes a housing 12 formed by adjacent cup-shaped members 13and 14 secured together at their periphery by bolts 15 and defining aninternal cavity 16. A drive shaft 18 is supported by hearing units 19and 20 to rotate about its longitudinal axis 22. A cage 24 secured tothe shaft is formed by a web 25 and an annular shoe ring 26 held theretoby engaging shoulders 27 and bolts 28. A plurality of inwardly facingshoe surfaces 29 are disposed on ring 26 symmetrically about the centeraxis 22. The cage 24 and shoe surfaces 29 rotate with shaft 18 about thelongitudinal center axis 22 which can also be considered as the centeraxis of the unit 10.

Annular end plate 30 is received over spacer ring 31 on the shaft 18between nut 32 and the inner race of bearing unit 19 and is secured tohousing 12 by bolts 33. Seals 34 and 35 between the moving shaft, spacerand housing prevent fluid leakage from the cavity 16.

A pintle element 36 (FIG. 3) is supported within bore 37 of the housing12 by spaced bearing units 38 and 39 to rotate about its longitudinalaxis 40. Axis 40 is parallel to but offset from the longitudinal axis 22of the shaft 18. Conventional means including a worm gear 41 keyed topintle 36 and engaged by a driving pinion 43 (shown in phantom) operateto rotate the pintle. O-ring 45 between gear 41 and housing 12 seal thebore 37 as required. The pintle 36 has a cylindrical crank portion 42disposed about a longitudinal center axis 44 parallel to but offset fromthe pintle axis 40 by the same offset as between the pintle axis 40 andthe shaft axis 22.

Cylindrical member 46 has a hub 47 defining a throughbore 48 whichmatably receives crank 42. The cylinder member 46 has a plurality ofradially extending projections 49 each having an internal cylindricalopening defining a cylinder 50. An opening 51 extends between theperiphery of through-bore 48 and the cylinder and is approximatelyone-half the area of the cylinder. A cylindrical piston 54 (FIG. 10) isreceived matably within each cylinder 50 and defines therewith a fluidchamber 55 communicating with the openings 51. The piston 54 has anoutwardly disposed bearing surface 56 matable with the shoe surface 29on the cage 24. Light compression springs 57 hold the pistons 54 againstthe cage 24 when the unit is not operating.

The crank 42 (FIG. 3) includes opposing slots or ports 58 and 60radially aligned with the through openings 51 in the cylinder member 46.The ports 58 and 60 are separated from one another by a rib 61 extendingacross the crank 42 and terminating at the outer periphery thereof onspaced lands 62 each of a given lap. A strengthening rib 63 extendstransversely of the ports 58 and 60 to a position short of the peripheryof the crank. Thus rotation of the cylinder member 46 on the crank 42exposes each opening 51 alternately to the ports 58 and 60.

The pintle 36 has spaced bores 64 and 66 extending axially thereof andcommunicating respectively with the opposing ports 58 and 60. Bores 64communicate through radial passage 65 with annular recess 67, and bores66 communicate through radial passages 69 with annular recess 68. Therecesses 67 and 68 are sealed from each other by O-ring gaskets 70received in grooves 71 and engaging the periphery of bore 37. Radialthreaded taps 72 and 73 in housing 12 communicate respectively withannular recesses 67 and 68 for separate connection to the high pressureand low pressure fluid sources, as is I well known in the art.

The subject hydraulic unit 10 operates in a mannner that eliminates thereaction couples between adjacent members caused by transformationbetween fluid pressure in the chambers 55 and rotation of the shaft 18.The cage 24 and crank 42 form the reaction elements opposing the piston54 and cylinder 50 (within cylinder member 46) on opposite sides ofchamber 55. The cylinder member 46 rotates freely on the crank 42 at thesame rate as the shaft 18 and cage 24 through the interactions of themating of pistons 54 on the cage 24. The cylinder member 46 followsfreely since it acts only to confine the fluid pressures in chambers 55and to communicate the chambers alternatively through the ports 58 and60 with the high and low pressure fluid sources.

As noted above, the shaft 18, the cage 24, and the spaced shoe surfaces29 all have a common center axis 22 which is parallel to but offset bysome eccentricity from the center axis 44 of the crank 42. During onerevolution of shaft 18 and cage 24, and thus of the cylinder member 46,the distance normal to each shoe surface 29 between said shoe surfaceand the crank center axis 44 varies between a maximum and a minimum.This variation in radial distance causes the reciprocation of thepistons 54 within the cylinders 50. It is apparent that if a plane wereextended through the spaced center axes 22 and 44, the top dead centersof the pistons 54 in the cylinders 50 would occur on that plane, themaximum and minimum being apart. Accompanying this radial reciprocationrelative to the cylinder 50, each piston 54 translates to each side ofits top dead center position relative to shoe surface 29 along the shoesurface a distance equal to the eccentricity of the unit.

The fluid pressure confined within each chamber 55 acts along thelongitudinal center axis of cylinder 50, which axis extends through thecenter axis 44 of crank 42 and the centroid of piston bearing surface56. At all positions other than the top dead center positions, thisaxial force is displaced from the center axis 22 of the shaft 18 andcage 24 to exert a turning moment or torque on the cage about its axis.The cage is one of the reaction elements and is also directly connectedto the shaft for common related movement therewith. The cylinder member46 follows the rotation of cage 24 freely and without resistance otherthan minor film friction losses. This offset application of the axialforce directly on the moving reaction member, that is the driving ordriven cage 24 connected to shaft 18, transforms between the fluidpressure forces and mechanical shaft torques. Furthermore, regardless ofwhether the unit is working as a pump or motor, or whether the shaft isrotating in one direction or the other, the transformation occursdirectly, without reaction couples between the members caused by thefluid pressure.

It will be apparent that a cocking force of minor magnitude will bepresent, caused by fluid friction in the films between the adjacentmembers. However, the magnitude of this force and the resulting coupleis negligible compared to the couple equal to the full shaft torquecaused by operating fluid pressures; a couple which this inventioneliminates.

FIGS. 7 and 8 show the fluid film bearing 20 in greater detail. Thebearing consists of a hub 80 connected centrally at circumferentiallyspaced portions by ribs or spokes 81 to an outer rigid rim 82. The rim82 is received within annular recess 83 of the housing 12 and heldtherein by a plurality of circumferentially spaced bolts 84 extendingthrough apertures 85. The hub 80 has cylindrical centerbore 86 from.0005 to .002 of an inch larger than the cylindrical shaft 18 receivedtherein, depending on shaft diameter. Thus there is radial clearance ofapproximately .00025 to .001 of an inch symmetrically around the shaft.

The inner periphery of the bore 86 has thereon a plurality of uniformlyspaced recesses 87. The bearing shown is only by way of exemplificationin which six equally spaced recesses 87 are disposed around the bore 86,each recess 87 being separated by axial land areas 88 and bycircumferential outer land areas 89 and an intermediate land area 90.The bearing shown has two such axially spaced bearing sections separatedfrom one another by the circumferential land area 90.

Each recess 87 has a radial bore 92 communicating therewith and with asource of high pressure hydraulic fluid confined within annulardistributing tubes 93. Manifold 94 connected to tubes 93 communicateswith tap 96 through flexible tubing 95. The tap 96 is maintained underhigh fluid pressure comparable to that of the unit 10, and can in factbe taken from the same source by a double check valve (not shown)connected, for example, between tap 96 and both taps 72 and 73. Arestrictor 97 of fixed size orifice is secured in each bore 92.

The basic operation of the subject film bearing 20 is similar to thosecommonly known as hydrostatic bearings. The high pressure source offluid at tap 96 continually presents high pressure through the radialpassage 92 and the restrictor 97 to the circumferentially spacedrecesses 87. When shaft 18 is centered in bore 86, the clearances fromeach recess between axial land areas 88 and circumferential land areas89 and 90 to the adjacent recess or to cavity 16 are all uniform and,therefore, the fluid pressures in the recesses 87 are maintainedgenerally constant.

When, however, a radial load is applied to shaft 18 to causedisplacement of the shaft within the bore 86, the clearances between theconfining land areas and the shaft in the direction opposing the loadare reduced while the clearances on the remote side of the shaft areincreased. The variation in clearance between the land areas surroundingeach recess 87 causes a variation of fluid resistance from the recess tochange correspondingly the fluid flow and fluid pressure. The restrictor97 has such fluid resistance that through flow at normally balancedconditions causes a pressure drop across the restrictor of approximatelyone-half or one-third of the high pressure source at tap 96. Theincreased fluid flow from the remote recess, because of the increaseclearance between the confining land areas and the shaft, causesrestrictor 97 in the passageway 92 to have a greater effect on thepressure drop of the fluid delivered to the recess. Thus on a remoteside of the shaft, the fluid pressure of the recess 87 is materiallyreduced from that of the balanced film condition.

Conversely, the displacement of shaft 18 within bore 86 causes areduction in clearance between the land areas opposing the load and theshaft to increase the flow resistance. Thus through-flow decreases sincethe total resistance to flow increases. At reduced flows, the throttlingeffect of restrictor 97 on the fluid pressure admitted to the recess 87is reduced to increase the fluid pressure within the recess.

Thus the increased and decreased radial clearance between the pressureconfining land areas of the recesses 87 in line with the shaftdisplacement and thus the load, causes the fluid pressures within therecesses to change. The differential fluid pressures on the oppositesides of the shaft tend to center the shaft within the bore untilbalance with the applied load is attained.

Similarly, axial misalignment of shaft 18 can be adequately absorbed bythe fluid pressure differences established in the axially spacedrecesses defined by land areas 89 and 90. Thus when there is a force onshaft 18 tending to tilt the shaft in a plane extending through itslongitudinal axis, the differential fluid pressures act within each ofthe recesses on the opposite sides of the shaft and at oppositelongitudinal ends of the bore 86 to produce a counteracting force; thusavoiding direct metal-to-metal contact.

The counteracting force of the fluid film tilts the hub somewhat tofollow the shaft, the tilting being absorbed by flexure of the radialspokes or ribs 81. The transverse cross-section of each rib 81 hassuflicient strength to support adequately any radial load acting on thebearing 20 to maintain the centerline of the bearing at the same generaltransverse position. The relatively high slenderness ratio, or the ratioof length of rib 81 compared to its transverse cross-section givesflexure to the rib to permit limited angular deflection of hub 80 withrespect to rim 82. Thus, any axial misalignment of the shaft 18 can beabsorbed by changes in the film thickness and resulting pressuredifference, and by flexure of the radial ribs 81.

The load from fluid pressures confined in chambers 55 between crank 42and the rotating cage 24 supported by shaft 18, produces a large bendingmoment and radial load on the projecting cantilever end of the shaft.The bearing 19 and the herein described fluid film bearing 20 supportand maintain axially spaced points on the shaft 18 in generally fixedtransverse positions. The high radial load, plus the moment of cage 24on shaft 18, causes the shaft to bend in a curved manner generallybetween the confining spaced bearings 19 and 20. The misalignment of theshaft caused by the above deflection can now be adequately absorbed bythe self-aligning film bearing 20 without metal-to-metal contact.

The resistance torque, or loss, of the bearing 20 is small because theforce required to shear an oil film is a function of the area of thefilm and is inversely proportional to the thickness of the film; in thisdesign a thin film .exists only at the land surfaces, which are small inarea.

Moreover, the bearing is smaller in diameter than existing anti-frictionbearings of long life and good reliability and hence the loss isappreciably less.

The fluid pressures confined within each chamber 55 also act in part onthe cylinder member 46 surrounding the opening 51, and are in partcommunicated through the radial openings 51 to the ports 58 and 60. Thefluid in the ports 58 and 6t) acts against the portion of the cylindermember 46 other than opening 51 in line with the ports, and leaks frombetween the crank 42 and bore 48 in the form of a fluid film. When theunit is operating, one of the ports will be under greatly higher fluidpressure than the other. Although the effects on cylinder member 46 offluid pressures within the chambers 55 and ports 58 and 60 generallyoppose each other, there still results a vector component forcegenerally in a direction transverse to the rib 61 connecting the lands62. Unless this phenomenon (commonly called separating force) iscorrected, it causes the mating crank 42 and cylinder member 46 on thehigh pressure side to separate, thereby tending to bind the low pressureside in metal-tometal contact.

FIGS. 3, 5 and 6 show a particular embodiment of means operable toovercome the separating force, thus eliminating metal-to-metal contact.Fine circumferential grooves and 106 in crank 42 are formed adjacent theports 58 and 60 respectively and extending generally parallel thereto.The grooves 105 and 106 are separated from each other by lands 107, andeach from its adjacent port 58 or 60 by a narrow land area 108 generallybetween A to A of an inch across, depending on the diameter and lengthof the bore and radial clearance, etc. The grooves themselves areapproximately 4 to A of an inch across and only a few thousandths of aninch deep. Land area 109 separates the grooves from the outside edges ofcylinder member 46 mating on the crank 42.

FIG. shows a passage 112 that intercommunicates axial bore 64 and groove106 on one side of crank 42 while a passage 113 intercommunicates bore66 and groove 105 on the opposite side of the crank. A restrictor 115 ofhigh fluid resistance is disposed in each passage 112 and 113 and isoperable to throttle the through-pressure by approximately one-third orone-half when exposed to balanced flow conditions.

Referring now to FIG. 6, the pressure forces and distributions onopposite sides of crank 42 (ignoring the effect of openings 51) areshown by the appropriate areas, generally designated 116. Assume thatport 58 and passages 64 are exposed to the high pressure fluid, whileport 60 and passages 66 are exposed to the low pressure fluid. Thus, theseparating force caused by the previous mentioned differentials inpressure acting on cylinder member 46 would ordinarily tend to causebinding of the right side of crank 42 (FIG. 5) against the cylindermember. To counteract this tendency, the high pressure from bore 64 iscommunicated through passage 112 across restrictor 115 to groove 106.The film clearance is extremely small adjacent grooves 106 so the filmpressure quickly builds up. Clearance adjacent port 58 initially isquite large so that leakage of the high pressure fluid across land area108 to groove 105 reduces the pressure considerably, groove 105 beinginterconnected across passage 113 and restrictor 115 further throttlingthe pressure to port 60. The restrictors 115 materially reduce flowthrough the passages 112 and 113 so that actually little loss because offluid flow results. The pressures at the grooves 105 and 106 then aredissipated approximately linearly across land areas 109 to the end ofthe cylinder member on the crank or across land area 108 to low pressureport 60.

The integrals of opposing fluid pressures acting over the appropriatelychosen land and groove areas balance cylinder member 46 moresymmetrically on the crank 42 to eliminate any direct metal-to-metalcontact. The land and groove areas and their positioning on the crankfrom the ports can be accurately determined to counteract the separatingforce previously encountered.

It has been noted that each piston 54 has a generally larger bearingsurface 56 matable with the shoe surface 29 on cage 24. Each pistonfurther has a through-bore 117 (FIG. extending from the defined fluidchamber 55 to a recess 118 on the inner face of bearing surface 56separated from the periphery thereof by land area 119. A restrictor 120is fixed within the through-bore 117. As fluid pressure in chamber 55forces the piston 54 and thus bearing surface 56 toward shoe surface 29,the fluid is simultaneously delivered at a reduced pressure viathrough-bore 117 and the restrictor 120 to recess 118 on the inner faceof the bearing surface.

The outwardly acting radial forces on piston 54 will be balanced whenthe average fluid pressure acting within recess 118 and film pressure onland areas 119 produces a resultant opposing force of equal magnitude.Under any balanced condition, the clearance between the surfaces 29 and56, or the film thickness will be of a certain value, and thus of agiven fluid resistance.

If the film thickness should increase; because of reduced filmresistance, flow would increase, the restrictor 120 would have a greatereffect in throttling the fluid pressure to recess 118 between thesurfaces than with balanced film thickness to reduce the average filmpressure. The supporting capacity of the fluid film thereby is reducedwith the reduced film pressure, permitting a reduction in filmthickness, and automatically varying the film thickness and pressureuntil balance is established. Similarly, if the film thickness isreduced; the film flow would be reduced, so that restrictor 120 wouldhave less effect in throttling the pressure, and the pressure introducedto recess 118 increases to increase the average film pressure.

Thus the adjacent surfaces are supported spaced apart by a fluid filmregardless of any unbalancing tendency acting on the members. Therestrictor throttles the pressure to recess 118 as required to maintainbalance for all applied forces. It has been observed that if at balancedfilm conditions the restrictor 120 throttles the pressure by one-thirdto one-half of that of the pressure source, the balanced film thicknessdoes not vary appreciably with variation of chamber pressures, and thusof loads.

pressure.

As was mentioned above, the longitudinal center axis 40 of pintle 36 isoffset by a given distance from the longitudinal center axis 44 of crank42. Similarly, pintle 36 is supported in housing 12 with itslongitudinal center axis 40 offset a similar distance from thelongitudinal center axis 22 of the shaft 18, cage 24 and shoe surfaces29, or the center axis of the unit 10.

FIG. 9 shows an operational sketch, greatly enlarged but similar to thefront section view (FIG. 2) of the variable flow per cycle adjustmentfeature for the unit. The unit center axis of the shaft 18, cage 24 andthe shoe surface 29 is represented at 22; the center axis of pintle 36is represented at 40; and the center axis of crank 42 at maximumeccentricity is represented at 44. The eccentricity of the unit 10 isthe distance from crank center axis 44 to the unit center axis 22, as isshown for the maximum eccentricity along the maximum top dead centerline 22, 40 and 44. The stroke of pistons 54 in cylinders 50 is twicethe eccentricity, the maximum stroke being along top dead center line123, 22, 40 and 44 as indicated.

Since pintle 36 can rotate in housing 12 about its center axis 40, thelocus of points defined by the adjusted crank center axis 44 extendsalong line 124 in a circular path above pintle axis 40 and intersectsthe unit center axis 22. When pintle 36 is rotated by worm gear 41through an angle 126 crank axis 44 is shifted to 44a. Line- 128 betweenthe adjusted crank axis 44a and machine axis 22, or the adjusted topdead center of the eccentric, rotates through an angle 130 from themaximum top dead center line 22, 40 and 44. Through elementarytrigonometry it is apparent that angle 130 is one-half the angle 126.Thus for pintle rotation through any angle 126, the top dead center ofthe adjusted eccentric is shifted by angle 130 from the maximum top deadcenter along line 22, 40 and 44.

For an angle 126 of pintle rotation, the eccentricity is reduced fromits maximum along straight line 22, 40 and 44, to its adjusted valuedesignated by line 128 between 22 and 44a. The reduction of theeccentricity is shown by distance 132. Thus, as the shaft 18, cage 24and shoe surfaces 29 rotate about the unit center axis 22, thecontraction and expansion of the pistons 54 in the cylinders 50 variesby twice the adjusted eccentric distance 22, 44a, or by the adjustedstroke 134, 22, 44a. The eccentricity varies as a harmonic function fromits maximum 22, 40, 44 to zero at 22 as the pintle 36 is rotated fromthe top dead center position through angle 126 equal to 180; in fact asthe cosine of one-half the angle 126 of pintle rotation.

The crank 42 is rotated through a similar angle as the pintle 36. Thusas pintle 36 is rotated an angle 126, the land areas 62 of crank 42 aresimilarly rotated an angle equal to 126, as represented in FIG. 9 bypoints 62a defined by the intersection of line extending from machinecenter 22 at an angle equal to 126 (shown as 130 and 136), and theadjusted eccentric circle 137.

Thus, radial openings 51 communicating between fluid chambers 55 and theports 58 and 60, do not pass the lands 62 represented by the points 62auntil cylinder member 46 is rotated relative to crank 42 an additionalThus at balanced conditions of the film, the film thickness is generallyindependent of the operating angle 136 past the adjusted top dead centeraxis 134, 22 and 44. From points 138, across 44a and 134 respectively,to point 62a, represented by an angular rotation of cylinder member 46on crank 42 through angles 130 and 136 equal to pintle rotation 126, thepistons 54 are reciprocated within the cylinders 50 a distance 140without actually causing fluid flow through the unit 10. The fluid ineach chamber 55 is caused to reverse normal flow during relativerotation of cylinder member 46 on crank 42 through an angle representedby 130; but is permitted to resume normal flow during the subsequentequal annular rotation represented by 136.

Thus the effective adjusted stroke of pistons 54 in cylinders 50 is nottwice that of the adjusted eccentricity 22, 44a; but is that portionwhich is represented by line 142, 22, 142 on the adjusted top deadcenter line 134, 22 and 44a. The effective adjusted stroke varies fromthe adjusted stroke as the cosine function of one-half the angle 126 ofpintle rotation from to 180. Since the adjusted stroke also varies asthe cosine function of onehalf of angle 126, the effective adjusted flowper cycle, consequently varies from maximum to zero as the square of thecosine function of one-half the angle 126 of pintle rotation from 0through 180.

Thus, for pintle adjustment through an angle 126, the adjusted effectiveflow of unit is varied both by the double mounted eccentric shaft 18,and pintle 36 and crank 42, and by phasing lands 62 with respect to thetop dead center positions. A graph showing the flow capacity per cycleas a function of adjusted pintle rotation from 0 to 180 traces a reverseS (it being the square curve of the first quarter cycle of a cosinefunction) having generally curved opposite ends and having a generallylinear intermediate portion. This is quite different from similar graphsof conventional variable flow units wherein the adjustment is affectedeither singularly by varying the stroke or singularly by phasing theport lands with respect to the top dead center positions.

It will also be noted that the structure including the double eccentricrelationship of the shaft 18, pintle 36 and crank 42 although easilyfabricated is capable of withstanding rigorous service conditions. Sincepintle 36 does not rotate rapidly, bearings 38 and 39 need not have thecharacteristics required of film bearing 20 for rapidly rotating shaft18. A simple worm gear 41 keyed to pintle 36 can be used to adjust thecrank 42 with respect to shaft 18 to vary the flow capacity, aspreviously described. The mounting of the pintle 36 is such that anyadjusted position of the pintle can be adequately and accuratelymaintained by the Worm gear 41 and its driving pinion 43; although otherstructure equally as simple could be used.

While a single embodiment has been shown, it is obvious that variationstherein can be made which employ the basic concepts of the subjectinvention. Accordingly, it is desired that the invention be limited onlyby the claims hereinafter following.

We claim:

1. For use in a hydraulic pump or motor unit having a shaft adapted torotate at a high speed and to support a transverse load, an improvedfluid bearing, comprising in combination a rigid hub member having acylindrical bore adapted to snugly receive the shaft, said hub having onits bore periphery at least a pair of axially spaced circumferentialland areas interconnected by a plurality of circumferentially spacedland areas and defining thereby a plurality of separate spaced recessesof generally shallow depth, means including a source of fluid under highpressure and separate fl-ow passageways therefrom to each of the definedseparate recesses effective to communicate fluid under pressure to eachrecess to be confined therein from leakage between the land areas andthe adjacent periphery of the shaft by a fluid film of generally highfluid resistance, means including a restrictor in each of thepassageways of flow resistance comparable to that of the fluid filmunder normal film flow and clearance conditions, and means to supportthe hub relative to the unit including rib structure extending radiallyfrom the hub adapted to be secured at its outer portion to the unit andhaving a cross-section of high slenderness ratio and strength to resistgenerally any radial deflection of the hub but to accommodate generallysome tilting deflection of the hub relative to the unit.

2. For use in a hydraulic pump or motor unit having a shaft adapted torotate at a high speed and to support a transverse load, an improvedfluid bearing, comprising in combination a rigid hub member having acylindrical bore adapted to snugly receive the shaft, said hub having onits bore periphery a plurality of separate axially and circumferentiallyspaced recesses of shallow depth defining thereby at least three axiallyspaced circumferential land areas interconnected by a plurality ofseparate circumferentially spaced axial land areas, means including asource of fluid under high pressure and separate flow passagewaystherefrom to each of the defined separate recesses effective tocommunicate fluid under pressure to each recess to be confined thereinfrom leakage between the land areas and the adjacent periphery of theshaft by a fluid film of generally high flow resistance, means includinga restrictor in each of the passageways of flow resistance comparable tothat of the fluid film under normal film flow and clearance conditions,and means to support the hub relative to the unit, said supporting meansincluding rigid structure spaced radially from the hub adapted to besecured to the hydraulic unit, and rib structure extending radiallybetween and structurally interconnecting the hub and rigid structure andhaving sufficient cross-sectional strength to resist generally anyradial deflection of the hub relative to the rigid structure but havinga generally high slenderness ratio to accommodate generally some tiltingdeflection of the hub relative to the rigid structure.

3. For use in a hydraulic pump or motor unit having a shaft adapted torotate at a high speed and to support a transverse load, an improvedfluid bearing, comprising in combination a rigid hub member having acylindrical bore slightly larger than the shaft diameter adapted tosnugly receive the shaft, said hub having on its bore periphery aplurality of separate axially and circumferentially spaced recesses ofgenerally shallow depth defining thereby a pair of circumferential landareas adjacent the ends of the bore and at least one intermediatecircumferential land area therebetween, and defining a plurality ofcircumferentially spaced separate axial land areas interconnecting therespective circumferential land areas, means including a source of fluidunder high pressure and separate flow passageways therefrom to each ofthe defined separate recesses effective to communicate fluid underpressure to each recess, the fluid being confined in the recess fromleakage between the land areas and the adjacent periphery of the shaftby establishing a fluid film of generally high flow resistance, meansincluding a restrictor in each of the passageways of flow resistancecomparable to that of the fluid film under normal film flow andclearance conditions adjacent the respective recess effective tothrottle the fluid pressure admitted to the recess by one-third toone-half that of the high pressure source, and means to support the hubgenerally fixed radially of the center axis of the shaft but generallytiltable relative thereto so that the longitudinal center axis of thehub can be tilted relative to the normal longitudinal center axis of theshaft while having the axes intersect somewhere generally within theaxial dimensions of the hub, said supporting means including an annularrigid structure spaced radially from the hub adapted to be secured tothe hydraulic unit, and rib structure extending radially between andstructurally interconnecting the hub and annular structure and havingsufficient cross-sectional strength to resist generally any radialdeflection of the hub relative to the annular structure but having agenerally high slenderness ratio to accommodate generally some tiltingdeflection of the hub relative to the annular structure.

4. A hydraulic pump or motor unit, comprising a housing, an annular cagehaving spaced shoe surfaces disposed symmetrically of its center axis, adrive shaft connected securely to the cage symmetrically of the centeraxis operable to support said cage and to be in direct drivingrelationship therewith, a pintle supported by the housing on arotational axis offset from said center axis, said pintle includingintegrally therewith a cylindrical crank having its longitudinal centeraxis offset from the rotational axis and disposed radially in line withthe shoe surfaces, a cylinder member having a through-bore matablyreceiving the crank, said cylinder member also having a plurality ofcylinders extending radially from the crank open toward the respectiveshoe surfaces, a piston member matably received in each of the cylindersand defining therewith a variable volume fluid chamber, each pistonmember being biased by fluid pressure in its chamber radially againstthe respective shoe surface so that the cage and the cylinder member andpistons are caused to rotate at the same speed and in the same directionabout their respective centers whereupon the distance between each ofthe shoe surfaces and the crank varies during each cycle by twice theeccentricity of the cage axis and the crank axis, means to support thecylinder member and the crank on fluid films of suflicient thickness toavoid direct bearing contact and including separate grooves at theinterface of the through-bore periphery and the crank, said supportmeans also including a passage of flow resistance comparable to thatcaused by the fluid film under normal film flow and thickness conditionsoperable to supply each of the groove-s with fluid at pressuresgenerally inversely proportional to the radial clearance adjacent therespective groove, means to rotate the pintle about its rotational axisto adjust the eccentricity of the cage axis from the crank axis operableto vary the flow per cycle of the unit and self-aligning bearing meansfor supporting the shaft and the cage said bearing means 5 including arigid hub having a bore closely surrounding the shaft adjacent the cagemeans including a source of fluid under high pressure and flow passagesthereform to the interface of the bore operable to establish a fluidfilm between the hub and the shaft each of said passages having aresistance to flow comparable to the flow resistance of the fluid filmitself under normal film clearance conditions and flexible structurebetween the hub and housing supporting the former so as to accommodate alimited tilt thereof relative to the normal longitudinal center axis ofthe shaft.

No references cited LAURENCE V. EFNER, Primary Examiner.

UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION Patent N00 3,211,105 October 12, 1965 Vannevar Bush et alt Column 12, line 11, after"unit" insert a comma; lines 12 and 14, after "cage", each occurrence,insert a comma; line 15, for "thereform" read therefrom line 17, after"shaft" insert a comma; line 20, after "conditions" insert a commaflSigned and sealed this 28th day of June 1966,

(SEAL) Attest:

ERNEST W. SWIDER EDWARD J. BRENNER Lttesting Officer Commissioner ofPatents

1. FOR USE IN A HYDRAULIC PUMP OR MOTOR UNIT HAVING A SHAFT ADAPTED TOROTATE AT A HIGH SPEED AND TO SUPPORT A TRANSVERSE LOAD, AN IMPROVEDFLUID BEARING, COMPRISING IN COMBINATION A RIGID HUB MEMBER HAVING ACYLINDRICAL BORE ADAPTED TO SNUGLY RECEIVE THE SHAFT, SAID HUB HAVING ONITS BORE PERIPHERY AT LEAST A PAIR OF AXIALLY SPACED CIRCUMFERENTIALLAND AREAS INTERCONNECTED BY A PLURALITY OF CIRCUMFERENTIALLY SPACEDLAND AREAS AND DEFINING THEREBY A PLURALITY OF SEPARATE SPEED RECESSESOF GENERALLY SHALLOW DEPTH, MEANS INCLUDING A SOURCE OF FLUID UNDER HIGHPRESSURE AND SEPARATE FLOW PASSAGEWAYS THEREFROM TO EACH OF THE DEFINEDSEPARATE RECESSES EFFECTIVE TO COMMUNICATE FLUID UNDER PRESSURE TO EACHRECESS TO BE CONFINED THEREIN FROM LEAKAGE BETWEEN THE LAND AREAS ANDTHE ADJACENT PERIPHERY OF THE SHAFT BY A FLUID FILM OF GENERALLY HIGHFLUID RESISTANCE, MEANS INCLUDING A RESTRICTOR IN EACH OF THEPASSAGEWAYS OF FLOW RESISTANCE COMPARABLE TO THAT OF THE FLUID FILMUNDER NORMAL FILM FLOW AND CLEARANCE CONDITIONS, AND MEANS TO SUPPORTTHE HUB RELATIVE TO THE UNIT INCLUDING RIB STRUCTURE EXTENDING RADIALLYFROM THE HUB ADAPTED TO BE SECURED AT ITS OUTER PORTION TO THE UNIT ANDHAVING A CROSS-SECTION OF HIGH SLENDERNESS RATIO AND STRENGTH TO RESISTGENERALLY ANY RADIAL DEFLECTION OF THE HUB BUT TO ACCOMMODATE GENERALLYSOME TILTING DEFLECTION OF THE HUB RELATIVE TO THE UNIT.